Method for a controlled supply of lubricant to an antifriction bearing

ABSTRACT

A lubrication system for supplying lubricant to a plurality of antifriction bearings composed of inner and outer bearing rings and a plurality of rolling elements mounted between the rings. The lubricant is intermittently supplied to the region of the rolling elements, with the amount of each intermittent supply being determined as a function of operating data which are continuouslu acquired from the bearing.

This application is a divisional of application Ser. No. 08/958,031,filed Oct. 27, 1997 now U.S. Pat. No. 5,971,107, which in turn is adivisional of application Ser. No. 08/325,298, filed Feb. 28, 1995 nowU.S. Pat. No. 5,711,615.

BACKGROUND OF THE INVENTION

The present invention relates to antifriction bearings having alubricant supply system wherein one of the bearing rings is providedwith a lubricant supply duct. Antifriction bearings of this type areknown from "Konstruktion von Spindel-Lager-Systemen fur dieHochgeschwindigkeits-Materialbearbeitung," Manfred Weck et al., expertVerlag, Ehningen.

In these antifriction bearings, a cleaning of the antifriction bearingby flushing occurs along with the lubricant supply. As a result, thereis always more lubricant supplied into the antifriction bearing thanneeded. It is therefore necessary to remove and discharge the excessivelubricant. Also, additional work energy must be applied to compensatefor "splash" work loss, and also an additional heating of the bearingoccurs as a consequence of the splash work, as well as a wear oflubricant.

EP 350 734 also discloses devices for supplying a lubricant toantifriction bearings. In these antifriction bearings, the lubricant issupplied through conduits, or nozzles, through which the lubricant mixedwith an air current is blown into the bearing.

A disadvantage of this embodiment is that besides the supply oflubricant to the point of lubrication, the environment is contaminatedby the lubricant. Moreover, this type of construction causes arelatively high lubricant consumption, since besides the lubricantrequired for the actual point of lubrication, it is also necessary toconsider the loss of lubricant which is blown into the environment.

While the selection of a relatively highly viscous lubricant allowsthese losses to be relatively low, they cannot be entirely avoided. Onthe other hand, highly viscous lubricants for certain applications,interfere with the easy motion of a bearing, as is required inparticular for godets for yarns, which rotate at a very high speed andsubstantially free of resistance. In particular in the case of godets,the contamination of the environment by oil mist is of relevantimportance, since it is absolutely necessary to prevent the product frombeing soiled.

It is also known to provide such antifriction bearings with a permanentfilling or lubrication, the bearing in this instance being hermeticallysealed by means of washers arranged on both sides of the rollingelement.

Prerequisite for such a permanent lubrication are correspondingly pasty,or highly viscous lubricants, so as to ensure a long-term sealing of theantifriction bearing. However, this in turn has the disadvantage thatthe viscosity of the lubricant is highly dependent on the temperature ofthe bearing.

It is therefore the object of the present invention to further developthe supply of lubricant to antifriction bearings, so as to ensure with alow lubricant throughput an effective and substantially lossfreelubrication of the antifriction bearing.

SUMMARY OF THE INVENTION

This object is accomplished by the provision of an antifriction bearingwhich comprises inner and outer bearing rings, an inner race formed onthe inner bearing ring and an outer race formed on the outer bearingring, and a plurality of rolling elements confined between the races.Also, a duct is provided in one of the bearing rings and which isconnected to a lubricant supply device, and the duct terminates at anoutlet opening in the region of the race of the respective bearing ring.

Contrary to all previously known technologies, the configuration of theantifriction bearing in accordance with this invention, allows theantifriction bearing to operate with a minimally possible and yetoptimal quantity of lubricant. The combination of these two influentialparameters for the service life of an antifriction bearing still leads,despite smallest quantities of lubricant, to an extended service life,and this with the least burden to the environment.

The invention also has the advantage that the lubricant is supplieddirectly and exclusively into the region of the rolling elements. It istherefore necessary to adjust the quantity of lubricant supplied only tothe immediate requirements of the antifriction bearing. The lubricantmay be supplied intermittently in small measured quantities, since thetotal quantity of the lubricant supplied is fed and used exclusively inthe region of the rolling elements.

Depending on the application, the lubricant may be supplied in a solid,pasty, or liquid form, i.e., in compact form without being mixed withair. The direct supply of the lubricant into the region of the rollingelements causes the lubricant to settle free of mist on the races of therolling elements, and so as to result in the buildup of a thin lubricantfilm between the rolling elements and their races.

Based on the small amount of lubricant required for such antifrictionbearings, it is preferred to proceed from the fact that a single duct inone of the bearing rings will suffice to supply the lubricant in therequired quantity. However, several ducts may be employed which arearranged, one following the other, in a single axial plane of theantifriction bearing or in the circumferential direction.

Since it is necessary that the metered quantities of lubricant be verysmall, especially when supplied to small antifriction bearings of fastrotating godet on textile machines, the diameter of the duct can be verysmall, i.e. it may be in the millimeter range or less (for example, 0.5mm).

Depending on the particular installation, the duct may be arrangedeither in the inner ring or outer ring of the bearing. In addition, theinner ring or outer ring of the bearing may be formed by a machineelement.

It is preferred that the duct be located in the stationary bearing ring,since this has the advantage of a simple connection to a lubricantsupply line.

To optimize the lubrication effect, one should consider the centrifugalforce which is effective on the lubricant. Consequently, the lubricantmay need to be supplied to the inner ring.

It is often possible however to provide the duct in the rotating bearingring. In this instance, it would be necessary to connect the lubricantsupply line to the duct, via a meshing channel or a rotating fluidcoupling.

With its characteristic features, the invention makes it possible tosupply the lubricant to the race of the rolling elements along ashortest is possible path. Thus, these features can result in a directsupply of lubricant to locations where the lubricant is needed.

It is also preferred that the lubricant duct be located in a zone of thebearing surface which carries a -educed load, and be structured suchthat it terminates in a peripheral zone of reduced surface pressure. Ineach instance, the rolling elements are prevented from falling into theduct, so that a smooth run can be realized. Consequently, the smooth runis likewise not disturbed by the duct. Thus, periodically occurringvibrations in the bearing are avoided, so as to permit the rollingelements to move in their races substantially evenly and free ofvibrations, and to further prevent a damage to the surfaces of therolling elements and their races by periodically occurring vibrations ofthe bearing.

Likewise, it is avoided that in the course of time the antifrictionbearings close the duct by their rolling motion.

A zone of a reduced load on the bearing may be created in anantifriction bearing by a special structuring of the race of the rollingelements. In this instance, the race of the rolling elements is designedsuch that the rolling element is supported primarily in two annular,spaced-apart regions. Located between these regions of support is anannular zone with a lesser pressure per unit area, in which the duct orducts can be arranged. Regardless of the type of construction, however,a zone of reduced load develops in each antifriction bearing on the sidefacing away from the transverse force of the bearing. The transverseforce operative on an antifriction bearing is the external forcedirected transversely to the shaft axis, which keeps the shaft inbalance together with the external load as well as the other transverseforces of the bearing. External loads and transverse forces of thebearing which occur on a shaft, lie more or less in an axial plane. Alsolocated in this axial plane is the duct, namely, on the side facing awayfrom the transverse force of the bearing.

In one embodiment, the duct is supplied with lubricant from the outside,via a groove which communicates with the duct and extends in the samenormal plane. In this embodiment, the angular position of thecorresponding bearing ring, when installed, is of no significance, sincethe lubricant groove is constructed as an annular groove, which extendsat least over a partial circumference of the bearing seat, and istherefore connected with the lubricant supply line in all angularpositions when installed.

In particular for high pressures of the lubricant, it may be appropriateto provide annular seals on both sides of the lubricant groove, so as topermit the formation of an impermeable annular space in the region ofthe lubricant groove, which is opened only for supplying the lubricantinto the antifriction bearing through the duct.

The annular seals may be arranged in annular ring grooves formed eitherin the bearing ring or in the housing, or in the bearing seat on theshaft side for the bearing ring, or even in a combination of these twopossibilities.

The duct is preferably connected to the pressure side of a lubricantpump, which serves to supply the lubricant under a predeterminedaccurate pressure. In this manner, when viewed over a period of time, analways uniform quantity of lubricant is supplied.

It is likewise possible to realize this advantage by connecting the ductwith a lubricant reservoir which is maintained under pressure. In thisinstance, it will suffice, because of the low consumption of lubricant,to load the pressure reservoir only from time to time, for example bybiasing a compression spring, which is operative on a piston that forcesthe lubricant out. In the meantime, i.e., while the pressure in thereservoir is slowly released, a pressure drop is accepted. The pressurereservoir thus need not be operated in a predetermined pressure range,and it need not be constantly adjusted to a single pressure value.

The following developments are especially advantageous for supplyingseveral points of lubrication as they exist on multiposition textilemachines.

Since antifriction bearings of the present invention permit an operationwith a minimal, but still optimal quantity of lubricant, it is desirablethat the lubricant be supplied from a source of lubricant under pressurevia a metering device. Metering devices of a variety of types may beutilized, and they may include, for example, volumetric meteringdevices. When using the latter, a predetermined, small volume quantityis first collected and then displaced. The displacement in this instanceoccurs in time intervals, i.e., at a frequency which determines themetered quantity. The volume which is displaced each time, remainsconstant. However, the metering device may also be of flow-limitingtype, such as, for example, magnetic valves which are opened and closedat predetermined time intervals and for a predetermined period of time.In this instance, the metered quantity is determined on the one hand bythe opening times, and on the other hand by the throttle cross sectionsof the valve.

The invention is based on a very low consumption of the lubricant. Thisallows an unnecessary filling of the antifriction bearing to be avoided,which is not needed for the lubrication. At the same time, this avoidslosses by splashing, i.e., an energy consumption which results from thefriction of fluid, and causes likewise high bearing temperaturesespecially at high rotational speeds. Due to the low lubricantconsumption, it is possible to arrange sealing disks at both sides ofthe races for the rolling elements.

Since the invention proceeds basically from a very low consumption oflubricant, sealing disks arranged on both sides of the rolling elementraces may bring advantages. These characteristics allow to prevent inparticular an uncontrolled evaporation of the lubricant, for example,caused by a high bearing temperature. To a certain extent the sealingdisks may be slightly permeable, so as to prevent an excessive fillingof the bearing on the one hand, and to realize a dustproof sealing onthe other hand.

However, an excessive filling of the antifriction bearing should beavoided in any event, so as to prevent additional losses by splashing,in particular at high rotational speeds, and the attendant high bearingtemperatures.

The antifriction bearing of the present invention is particularlysuitable for supporting rotable godets on a textile machine, whichcomprises a plurality of processing stations, since it is now possibleto use fast rotating antifriction bearings with a long service life,while preventing environmental contamination, in particular, the soilingof the product by oil.

As a result, special importance is attached to the method aspects of thepresent invention and wherein the amount of the lubricant intermittentlysupplied to the antifricticn bearing is determined in a control unit, inwhich a basic adjustment established by predetermined operatingparameters is modified as a function of operating data or data ofconditions which are continuously acquired on the bearing.

While in principle such a method is known from the aforesaid EP 350 734,this method operates, however, with high losses. Although the lubricantis supplied in a very accurately metered quantity to a transporting aircurrent and supplied into the antifriction bearing, a portion of thevery accurately metered lubricant is again flushed out by the aircurrent. It is therefore necessary to supply a higher metered quantityof lubricant than is needed.

Only as a result of combining the measurement of parameters and thesupply of lubricant directly into the region of the rolling element racecan a consistency be expected to exist between the metered quantity oflubricant and the bearing-specific requirements for lubricant.

Therefore, this combination is especially suitable for use in theconstruction of textile machinery, for example, for supporting godets,where it is absolutely necessary to avoid a contamination by oil mist.

In accordance with the preferred embodiment of the invention, thelubricant is delivered to the individual antifriction bearings underincreased pressure, and supplied there in a metered quantity in highlycompact (fluid or pasty) form and free of mist. This offers thepossibility of a progressive lengthening of maintenance intervals ofsuch a multiposition textile machine. In this instance, it should benoted that such a textile machine comprises a plurality of bearingpoints, each of which is subjected to a certain wear.

The rotatably supported machine elements, such as, for example, godets,winding heads, are arranged along the direction of advance of a yarn,and they are contacted by the yarn until its takeup on a package. As aresult, there exists the problem of having to shut down the entiremachine, at least one processing station, when one of the antifrictionbearings is defective.

Thus, in each instance, the minimal service life of the bearingdetermines the length of an operating phase, during which the textilemachine operates continuously. Since upon the failure of one of theantifriction bearings all antifriction bearings are replaced for safetyreasons, it is of great interest to extend the service life of thebearings, even under the difficult operating conditions in the case oftextile machines.

Since this invention allows uniform and optimal operating conditions tobe provided for all antifriction bearings, it becomes possible toincrease the service life of the bearings at least substantiallyirrespective of the bearing load.

This is accomplished in that, despite of the plurality of bearingpositions, an individual adjustment of the lubricant supply to theactual lubricant requirement is realized for each antifriction bearing.

In other words, a constant correction of a basic adjustment occurs foreach individual bearing. The basic adjustment is predetermined at themanufacturer's end. It results from empirical values and is corrected bythe data of conditions of the individual bearing points. Among them arein particular the temperature. The data of conditions are input, forexample, in a control unit and compared with the data of the basicadjustment stored therein. The basic adjustment should correspond tooptimal operating conditions, so that the comparison of the data ofconditions with the basic adjustment allows to realize a practicallyideal supply of lubricant within the range of the optimal quantity oflubricant of each individual bearing.

Thus prerequisite for establishing the individual quantities oflubricant are the data of conditions which are individually obtained ateach bearing location. To this end, it is possible to generate, forexample, actual value signals from the bearing temperature, which areinput in a control unit. It is further possible to generate actual valuesignals from the shaft speeds occurring on each bearing, which are inputlikewise into the central control unit. Calculated from the input actualvalue signals is in each instance the optimal lubricant requirement,which is to be supplied to each individual bearing.

For purposes of monitoring larger bearings, it is possible to provide inadvantageous manner two or more temperature measuring points distributedover the circumference of the respective bearing, and to transmit to thecentral control unit the mean values which are determined from themeasured values.

It is possible, though, to obtain certain operating parameters forcontrolling the metering of the lubricant by determining performancegraphs. However, such performance graphs can always apply to only oneinstalled situation, since it is necessary to individually ascertain allparameters, such as rotational speed, heating temperature, bearingarrangement, etc., individually, and to functionally correlate them withthe metered quantity of lubricant.

In particular, in the case of textile machines, it has shown that thetemperature has a plurality of influential factors.

These influential factors do not directly and necessarily relate to theinfluential factors, which generate the bearing temperature by friction.In textile machines, the bearing temperature is falsified, for example,not only by the operating conditions of the machine elements (godet),but also by the length of operation and the state of wear of thebearing.

It has been recognized that heated godets, which rotate slowly, have ahigher bearing temperature than fast rotating godets. From this, it isobvious that the control/regulation of the metering of lubricant cannotproceed reliably by only measuring the bearing temperature.

It is also necessary to take into account that the relationship betweenbearing temperature and lubrication changes. This applies in particularfor the reason that wear has a separate influence on the bearingtemperature.

This gives rise to the further problem of finding an operatingparameter, which permits a lubrication of the bearing that is clearlyadapted to the need for lubricant. The term operating parameter orparameter of state is understood to be physical values which permit todescribe the actual state of the bearing.

In one embodiment of the invention, the conditions which are monitoredon the antifriction bearing for determining the metered quantity of thelubricant, include the vibration of the bearing, which is totallyindependent of the bearing temperature. The determinant "bearingvibration" is conceived to be an operating parameter or parameter ofconditions, which is no doubt indicative of the lubricating conditionsin the antifriction bearing. Thus, an actual state of the bearing isdetermined which is temperature-independent.

An advantage may be seen in that, while preventing assembly-specificperformance graphs for the progression of the bearing temperature,easily determinable operating parameters are available for controllingthe supply of lubricant in metered quantities.

A further advantage consists in that the condition of vibration isclearly indicative of the lubricant of the bearing, and thereforedetermines also the additional need as a result of wear. It is observed,for example, in a new bearing that a certain condition of lubricationcauses vibrations to a certain degree.

What results is a useful evaluation, in particular it is possible toevaluate the amplitude. Accordingly, for example, amplitude peaks areascertained, in that, for example, an upper limit is established for theamplitude, and that one determines the exceeding of the upper limit, theduration of the exceeding, the number of bearing vibrations, duringwhich the limit value is exceeded, or the repetitiveness of exceedingthe limit values.

It shows, however, that the lubricant requirements and the wear of thebearing can be determined synchronously, in particular in that certainfrequency ranges of the vibrations are determined, for example by theFourier analysis, and that their occurrence or the frequency of theiroccurrence are ascertained. A particularly indicative frequency rangelies between 200 kHz and 500 kHz. Likewise, this embodiment of theinvention allows to measure a certain vibrational behavior for eachantifriction bearing.

To determine the vibrational behavior, a vibration sensor is installedin the stationary portion of the antifriction bearing, the lubricationof which is to be metered. Bearing vibrations or oscillations occurringthere are picked up continuously or at certain time intervals. Dependingon the bearing load, rotational speed, and condition of the bearing,time intervals of hours are permissible. The occurring vibrations arethen analyzed. A simple method of analyzing consists in that certaintolerance limits are predetermined, that it is then determined whetherthe vibration amplitudes leave the tolerance range. In this case, atolerance band is established, within which the bearing vibrations areallowed. When the vibration amplitudes leave the tolerance band, as isthe case shortly before a dry operation sets in, it will be necessary torelubricate. To make sure that the amplitude of the bearing vibrationand the exceeding of the predetermined tolerance limits are anadequately accurate indicator of the individual lubricant need and/orthe wear, it may be recommendable to ascertain beforehand by test thelife cycle of a bearing test in its respective range of application.

A reliable indication can be obtained from the amplitude analysis, whenthe latter is applied to vibrations of a certain frequency range. Thevibrations occurring in the bearing range of an antifriction bearingrepresent a superimposition of vibrations of different frequency ranges.A plurality of these frequency ranges is not indicative of the state oflubrication and/or the wear. For example, it has shown that, withrespect to these criteria, vibrations in the frequency ranging from 200kHz to 500 kHz are typical. For this reason, is its suggested that allvibrations outside of this frequency range, which is known to beindicative, are preferably filtered out. The amplitude analysis is thenperformed only with the vibrations of the critical frequency range,i.e., for example between 200 kHz and 500 kHz. While it cannot beprecluded that also in the amplitude analysis, which relates only to acertain frequency range of the vibrations, an amplitude peak, whichleads to exceeding the intended frequency range, is based on asuperimposition of also such vibration frequencies, which are notindicative of the state of lubrication and wear, this method offershowever already an adequately accurate possibility of adapting veryaccurately the quantity of lubricant to the actual requirements oflubricant.

Moreover, it has also shown that it is possible to make the lubricantsupply more accurate. To this end, the determined vibrations areanalyzed (Fourier analysis). It has shown indeed that vibrations ofcertain frequency ranges will not occur in satisfactorily lubricatedbearings, it being presumed that satisfactorily lubricated bearings donot exhibit any noticeable wear. In any event, the amplitude height ofthe vibrations of this frequency range does not exceed a predeterminedmeasure. Upon occurrence of these vibrations, a metered, very smallquantity of lubricant will be supplied to the antifriction bearing. Themonitoring of the bearing vibration has shown that as a result thereofthe vibrations of this frequency range will again disappear. Should now,due to wear, the time intervals decrease between two successive statesof vibration (amplitude peaks, frequency ranges), it will be possible todraw therefrom a conclusion to the extent of wear. On the other hand, itis possible to predict that wear will lessen, when a characteristicstate of vibration recurs within a determined metering interval. In thisevent, it will be necessary to accordingly reduce the metering interval.Consequently, one can expect that the respective length of the meteringintervals allows to conclude the actual wear, so that it becomespossible to reliably predict the maintenance intervals of such amachine.

As a result it becomes possible not only to adapt the bearinglubrication very sensitively to the wear, but to make nonetheless alsoreliable statements as to the state of wear. Thus, for the first time, ametering of the lubricant occurs, which is dependent on both thelubricant requirements and the wear.

Operating parameters to consider include in particular bearingvibrations of selected frequency ranges and/or bearing vibrations havinga selected amplitude height, it being possible to measure the timeintervals between at least two successive, characteristic bearingvibrations or amplitudes. This allows to then determine the meteringinterval. A further improvement may be realized as follows: from thelength of the metering interval, one may determine a so-called shortenedmetering interval, the length of which is somewhat shorter than the timeinterval measured between the successive occurrence of a characteristicvibrational behavior. This allows to avoid that a lubricant deficiencyoccurs at all.

The essential aspect of this further development of the invention istherefore based on the combination of the occurrence of a characteristicvibrational behavior in the antifriction bearing due to dry operationand the inclusion of the thus obtained time intervals into a controlcircuit for a regulated/controlled subsequent metering of lubricant.

Since, in principle, a subsequent metering can proceed very rapidly, andsince on can basically presume a very fast distribution of the meteredlubricant into the race of the rolling elements, the time intervals ofthe subsequent metering of lubricant may correspond essentially to thetime intervals, at which the characteristic vibrational behavior after apreceding metered supply of lubricant is again detected.

A fact to refer to is that this method is considered not only for theantifriction bearings of the structure described herein, but alsowherever a controlled supply of lubricant occurs to an antifrictionbearing, in that the lubricant is supplied in a metered quantity as afunction of the progression of a predetermined operating parameter. Tothis end, the vibrations generated by the antifriction bearing aregathered, and the operating parameter is determined from a vibrationalbehavior which is characteristic of a lubricant deficiency.

Thus it is possible, for example, to meter the supply of lubricant bythe transport of the lubricant in the form of an oil mist, as isdisclosed, for example in EP 350 734, as well as in EP 26 488.

Preferred is to supply the lubricant to the individual bearings ingaseous or liquid form as lubricating oil and under an increasedpressure. To this end, suitable bores, if necessary also several boresfor each bearing, are provided preferably in locations of the bearingcarrying the least load. It is useful to provide in the outside of thebearing rings, or better however in the housing bores accommodating theouter rings of the bearing, annular channels connecting to thelubrication oil supply, through which the lubrication oil enters intothe bearings.

The measuring and delivery (metering) of the lubricating oil destinedfor the individual bearing can occur in different ways. Thus, eachbearing may be associated with its own pump, for example, a segment of amultiple pump (the segments of which, however, need to be individuallycontrollable with respect to the quantities to be delivered).Displacement pumps have shown to be especially suited. However, thelubricating oil may also be supplied locally by a valve, in particularcontrolled by a magnetic valve, from a pressure oil tank, which isrefilled by a pump from time to time.

In particular in these instances, it is possible to control the amountof lubricating oil respectively supplied to a bearing, while maintainingregular time intervals between the opening signals, by changing therespective pulse duration, or while maintaining the same pulse duration,by varying the time intervals between opening signals, which are adaptedto the momentarily required amount of lubricating oil.

In accordance with the invention, extremely small quantities oflubricant are required. Therefore, the problem of aging lubricant,sedimentation, resinification, or saponification may arise both in theinstance of a controlled lubricant supply to the antifriction bearing,and in the instance of an uncontrolled lubricant supply (for example,the supply of lubricant in a fixedly predetermined timing cycle).

The above problems may be alleviated by the provision of a closedcircuit supply line for the lubricant, which is connected with alubricant pump having a tank, and with a pressure increasing device,such as a non-return valve, for increasing the pressure of the tank.Also, between the pump and the non-return valve are tap lines and ametering device associated with each bearing.

This further development of the invention brings the advantage that theclosed-circuit supply system represents a constantly flushed main line,in which the lubricant is kept hermetically sealed against air, therebypreventing an artificial aging of the lubricant.

On the other hand, a certain quantity of lubricant is always incirculation. The constant circulation of the lubricant in theclosed-circuit supply system, allows to accomplish an approximatelycontinuous, automatic evacuation of air, and yet to supply to eachantifriction bearing always a very accurately metered, smallest amountof lubricant.

A throttle may be positioned in the closed circuit supply line upstreamof the non-return valve. The combination of a throttle with a subsequentnon-return valve, which opens toward the tank against a spring, ensuresthat despite the constant flushing of the main line an adequate pressureis always present for supplying a plurality of antifriction bearingswith their associated metering pumps or metering valves, as will bedescribed in greater detail with reference to the circuit diagram.

If, in the place of the subsequent non-return valve, a time-controlledstop valve is used as a device for increasing pressure, it will bepossible to realize a flushing of the closed-circuit supply systemindependent of its length and diameter, throttle dimensions, andtemperature.

The closed circuit supply line may be connected with a pressurereservoir, which is located in the forward flow portion of the supplyline. This construction takes account of the fact that the lubricant isan incompressible medium, so that at a corresponding frequency ofremoval, the stored pressure also decreases relatively fast. If theclosed-circuit supply system is provided with an additional pressurereservoir, the pressure in the closed-circuit supply will be madeuniform. Pressure pulsations are unable to develop by either by feedingthe oil by means of the pump, or removing the oil by means of themetering device. Furthermore, it is possible to supply the lubricantfrom time to time by means of a pressure pump, and to keep the pressurein this manner within a predetermined range during the lubricant removalcycles. To this end, it is necessary that the pressure reservoir keepthe oil under pressure. To do so, an oil reservoir may be used, the tankcapacity of which is biased by a spring-loaded or compressed-gas loadedpiston. Likewise, it is possible to subject the oil in the tank directlyto the load by a compressed gas.

An important characteristic is that the pressure reservoir should belocated in the forward flow portion of the closed-circuit supply system.This allows to realize a forced flushing in the direction toward theother end of the closed-circuit supply system, and to compensate in asimple manner for the pressure losses within the closed-circuit system.

The metering device may be a metering pump, such as a piston pump. Thisconstruction offers the advantage of an accurate measuring of eachmetered amount of lubricant. In particular, a piston-operated meteringpump in accordance with the invention delivers a fixedly predeterminedvolume of discharge. In this instance, it is only necessary to adjustthe startup frequency of the metering pump to the case of application orto control same as a function of the bearing condition.

In another embodiment, the metering device may be a valve, such as amagnetic valve. This has the advantage that keeping the valveperiodically open allows to occasionally flush the line between thevalve and antifriction bearing.

The tap lines preferably extend from the closed circuit line with adowngrade toward the metering device. This provides a favorable effecton the automatic bleeding of the tap line, since air not dissolved inthe lubricant will always attempt to remain in the closed-circuit supplysystem, whence is continuously discharged. This method, when viewed inthe course of time, has a favorable effect on a uniform supply oflubricant to the antifriction bearing. This advantage is realized bydownwardly inclined tap lines, in which any possibly present air alwaysrises upward as a result of buoyancy, so as to be entrained in theclosed-circuit supply system by the flushing operation.

The pressure in the closed circuit supply line is preferably maintainedbetween an upper limit valve and a lower limit valve, and the closingpressure of the non-return valve is between the upper and lower values.This feature may be used in particular for a fully automatic pressurecontrol, once the upper limit value as well as the closing pressure ofthe subsequent non-return valve are determined.

This allows to influence in addition the respective flushing time byadjusting the closing pressure of the non-return valve. If the closingpressure is near the upper limit value, the flushing time will beshorter than when the closing pressure is closer to the lower limitvalue.

This offers the advantage of an additional operational reliability whichis enabled alone by controlling the pressure.

In this connection, an additional advantage is achieved by having theinlet pressure of the metering device picked up by a pressure monitor,and such that when the inlet pressure falls below a lower limit valve, adisconnecting signal is emitted to shut down the associated bearing. Bythis arrangement, an operational reliability of the antifriction bearingmay be ensured, even when no lubricant is supplied. In this instance,the time delay is measured by the time which will lapse until thelubricant is used up in the antifriction bearing.

Advantageously, the lubricant is metered by hydraulic devices as furtherdescribed below. However, it should be expressly stated that thesemeasures are useful both in antifriction bearings without controlledmetering of the lubricant and in antifriction bearings with a controlledmetering of the lubricant.

Moreover these measures should expressly be viewed both in combinationwith a closed-circuit supply system as described herein, and alsoseparately therefrom.

Hydraulic devices for metering lubricating oils in antifriction bearingsare known, though, from U.S. Pat. Nos. 4,784,578 and 4,784,584. Theydisclose devices operated by a separate pump, in which a piston isactuated under the pressure of the lubricating oil delivered by thepump, and which pumps out the amount of lubricating oil being in thecylinder. These devices are structured in a complicated manner. Inparticular, they are provided with an inlet valve, which allows andeffects a filling of the cylinder as a function of the piston stroke.The filling is thereby subjected to inertia, i.e., it is dependent onthe movement of the inlet valve and the flow of the fluid. Consequently,the piston can be moved only so fast that adequate time is left for thefluid to follow.

When using such pumps in the field of a metered lubrication ofantifriction bearings, it is necessary to ensure a complete filling ofthe cylinder at a very fast piston movement. This requirement can berealized with a pump which comprises a cylinder, a pump piston arrangedin a guideway coaxial with the cylinder and adapted for movement in thecylinder by a power drive, an inlet valve connected to a fluid supply,and an outlet valve. Also, an inlet chamber communicates with thecylinder and the guideway, and the inlet valve is formed by the boundaryedge of the cylinder toward the inlet chamber, and the pump pistonincludes an end surface which is adapted to move between a position infront of the boundary edge and an immersed position in the cylinder.

The above solution is based on the fact that the inlet valve is formedby the piston itself and by the boundary edge between the cylinder andthe inlet chamber. However, it is situated at an unusual location,namely, the inlet opening is released practically only when -the pistonhas finished its intake motion. During this intake motion, the pistonhas produced in the cylinder a vacuum, into which the amount oflubricant to be metered flows in very swiftly, as soon as the pistonmoves with its front surface out of the emptied cylinder and into theinlet chamber.

In the further development of the invention the lubricant in the supplyline is put under increased pressure and retained by a controllablevalve, which has an outlet end in the lubricant supply line. Thelubricant is thus stored under a pressure, which is above theatmospheric pressure. This allows to accomplish that the lubricant canbe supplied in accurately metered quantities to the antifriction bearingat accurately scheduled times and during accurately predeterminedperiods of time. An essential difference from the metering pump lies inthe variable volume of lubricant per metering, since the opening time ofthe valve determines the supplied volume of lubricant. As a function ofthe respective viscosity, it is possible to realize the requiredpressure level in a very simple manner, also from the viewpoint ofreliability. Thus, low-viscosity lubricants will require likewise a lowpressure level, so as to enable a metered supply of lubricant in themeaning of this invention.

However, it should be expressly stated that also very highly viscouslubricants can be used. In this instance, it would be necessary to raisethe pressure to an accordingly high level, so as to be able to supply tothe antifriction bearing the required amount of lubricant during a shortopening time of the valve.

A further advantage lies in the metering device which is technicallyvery simple to realize, and needs to have only an externallycontrollable valve.

A still further advantage which results from the invention, lies in thatthe outflowing lubricant bears little on the environment, since theinvention, on principle, is based on very short opening times of thecontrollable valve. Thus, lubricant is supplied only in meteredquantity, which is needed by the antifriction bearing. This fact isbased on the recognition that the need of lubricant required by anantifriction bearing should be only small.

A certain throughput of lubricant through the antifriction bearing,however, has positive effects with respect to wear, since abrasiveparticles are carried away from the contact zones between rollingelements and their races by the lubricant, as it is throughput in thecourse of time.

The controllable valve in the lubricant supply line may be anelectrically controllable magnetic valve, which offers the advantagethat the valve can be activated rapidly both in direction of opening andin direction of closing, thereby making it possible to exactly maintainthe accuracy of the quantity of lubricant supplied in each instance alsoover a plurality of lubrication intervals.

The control of the quantity of lubricant supplied to each bearing mayoccur via a sequence of tripping signals of a predetermined duration andcontrolled frequency. It may also occur via a sequence of predeterminedfrequency and controlled duration. Each such control method offers thepossibility of guaranteeing with one and the same control system thesupply of lubricant for different antifriction bearings, installedsituations, rotational speeds and loads of the antifriction bearing, andabrasive wear, just to mention a few influential variables.

As a result, it is also possible to realize a modular structure of thecontrol system, which covers certain ranges of diverse requirements,while simplifying stockkeeping.

The lubricant supply line may be designed as a pressure reservoir, whichis preferably closed toward the inlet side by a non-return valve. Thisoffers the advantage that the lubricant supply lines, which are anyhowpresent, are capable of assuming the additional function of a pressurereservoir. It has been recognized by the invention that the lubricantsupply lines make available a certain volume, which allows to easilystore the lubricant, and this under an increased pressure. This type ofconstruction offers itself in particular for the retrofitting ofexisting antifriction bearings with an external supply of lubricant,since no further structural elements will be needed, except acontrollable valve and, if need be, a non-return valve on the inletside, as well as a device for increasing the pressure.

For certain applications, it is desirable that the lubricant be suppliedfrom a pressure reservoir which is biased by a pressure medium, such asgas. This applies, for example, when the existing volume of thelubricant supply line appears to be too small, so as to realize areliable lubrication of the antifriction bearing, when no lubricant isreplenished.

In this event, the pressure reservoir serves, on the one hand, to storethe excessive amount of lubricant, which cannot be accommodated in thelubricant supply line. On the other hand, it is possible to makeavailable in the pressure reservoir a space which is occupied by thepressure medium. The thus occupied volume of the pressure reservoir willcontinue to be available for the pressure medium, even when the devicefor increasing the pressure fails to function. This circumstance resultsin the advantage that such failures can be temporarily buffered.

It has been recognized that in most applications, there is no need toexactly maintain the lubricant pressure. As a consequence, this allows apressure control to be utilized which is adapted to requirements.

In so doing, it is possible to determine the upper or lower limit valuedirectly by a pressure measuring device. However, it is alsoconceivable, provided the need for lubricant is recognized, to realizethe control by a simple timing operation.

BRIEF DESCRIPTION OF THE DRAWINGS

In the following description, the invention is described in more detailwith reference to the drawings, in which:

FIG. 1 illustrates a first embodiment of the invention;

FIG. 2a is an axial front view of the bearing shown in FIG. 1 and FIG.2b is a side view of a pair of the bearings mounted on a common shaft;

FIG. 3 shows a further embodiment of the invention;

FIG. 4 shows an embodiment, which comprises a plurality of antifrictionbearings in a control circuit for a controlled supply of lubricant;

FIG. 5 shows an embodiment of the invention included in a closed-circuitsupply system;

FIG. 6 shows a possible embodiment of a metering pump;

FIG. 7 shows a further embodiment of a metering pump;

FIG. 8 shows an embodiment of the invention with a metering valve;

FIG. 8a shows an embodiment in accordance with FIG. 8 with aclosed-circuit supply system of FIG. 5;

FIG. 9 shows an embodiment of the invention with a metering valve and areservoir;

FIG. 10 is a schematic view of a spinning system for filament yarns withthe processing steps of extruding, drawing, and winding;

FIGS. 11a, 11b, and 11c illustrate the invention in the threeantifriction bearings with a no load zone in the race despite atransverse force; and

FIG. 12 illustrates the invention with an axially biased shaft bearingsupport.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Unless specified otherwise, the following description will apply alwaysto all Figures.

FIGS. 1-3, in particular, illustrate an antifriction bearing I with adevice 2 for an external supply of lubricant.

The structure of such an antifriction bearing comprises an outer ring20, an inner ring 21, and rolling elements 5 confined therebetween,which are normally secured in their position relative to one another bya cage 22, so that they roll equally spaced apart, along races 6a or 6i.Essential is that one of the bearing rings, in the illustratedembodiment outer ring 20, contains a passageway or duct 3 whichconnects, on the one hand, to a lubricant supply line 4, and terminates,on the other hand, in radial direction toward rolling elements 5 at 23.In the present embodiment, it thus matters that the duct 3 extendsthrough the bearing ring such that the lubricant leaving the outer ring20 through outlet end 23 is supplied directly to the rolling elements inthe region of their race.

To this end, it is possible, on the one hand, to arrange outlet end 23in direct vicinity of the zone contacted by the rolling elements. On theother hand, it will be advantageous to arrange outlet end 23 in acentral region of one of the rolling element races, in the presentembodiment, the outer race 6a. In this instance, the lubricant issupplied only where it is needed to reduce wear by building up asupportive lubricant film between the rolling elements and their race.

It is easy to visualize that duct 3 can also terminate in the inner race6i of the rolling elements, provided this is advantageous for theinstalled situation.

The race for the rolling elements is that portion of the antifrictionbearing rings, which is defined by radial planes which are contacted bythe rolling elements on both sides. It is that portion of the racemachined by grinding into the bearing rings, on which the rollingelements are allowed to roll.

Preferably, duct 3 extends in a central radial plane, when the bearingis axially biased. In this instance, a zone under little stress developsalways in the central radial plane of the bearing, so that in the courseof time the outlet end 23 of duct 3 is prevented from being closed bythe rolling motion. Furthermore, this type of construction has theadvantage that the position of duct 3 is independent of the respectivelyinstalled position, as will be described with reference to FIG. 12.

As is further shown in FIG. 1, the outer ring 20 is fitted into abearing seat 10. In the present embodiment, the bearing seat 10 isarranged in the surrounding housing. The end of duct 3 facing thebearing seat is cut in by a peripheral, annular lubricant groove 7. Thisannular lubricant groove 7 may be arranged in the outer ring 20 of thebearing. In the illustrated embodiment, however, the lubricant groove isarranged in the material of the housing, in a radial plane, in whichalso the duct 3 terminates on the side facing the bearing seat.

The lubricant groove 7 is again connected to lubricant supply line 4,through which lubricant is supplied to the antifriction bearing.

An advantageous further embodiment is shown in FIG. 1. In thisembodiment, annular grooves 8, 9 for receiving sealing rings 24, 25 arearranged on both sides of lubricant grooves 7. Such sealing ringsconsist of an elastic material, and are inserted into the correspondingannular groove 8, 9 with their diameter slightly projecting therefrom.When installing the outer ring 20 of the bearing, the sealing rings arecompressed by bearing seat 10 in radial direction, and thereafter restunder a bias between the bottom of their associated annular groove 8, 9and the bearing seat 10. In this manner, a reliable sealing is realizedon both sides of annular groove 7, so as to permit the lubricant to exitonly, as intended, through duct 3.

In the present embodiment, the annular grooves 8, 9 are arranged in theouter ring 20 of the bearing. When the annular grooves 8, 9 are arrangedin bearing seat 10, advantages in machining can be obtained, since thematerial of the bearing seat is easier to machine than the material ofthe bearing ring. Likewise, it is possible to arrange one of the annulargrooves in the bearing seat, whereas the other annular groove isarranged in the bearing ring.

As is further shown in FIGS. 2a and 2b, the antifriction bearing 1 isunder the influence of a force 11, which is applied to it, and which maybe the result of a force F of a yarn looping about a godet. Air which isalways present in the bearing will cause the inner ring 21 and the outerring 20 of the bearing to displace relative to one another in radialdirection (exaggerated in FIG. 2), so that a load zone 12 forms, inwhich the rolling elements are in constant engagement. Diametricallyopposite thereto, a substantially load-free zone 26 is formed, in whichthe rolling elements roll along their races 6a, 6i at most under aslight pressure. As can be noted, the outlet end 23 lies outside theload zone 12 between the rolling elements 5 and their races 6a or 6i. Itis recommended to arrange the outlet end 23 approximately diametricallyto the center of load zone 12.

As is further shown in FIG. 1, the lubricant supply line is connected tothe pressure side of a lubricant pump 13, which may be an intermittentlyoperated pump. Alternatively, the lubricant pump may operate extremelyslowly, and thus discharge continuously always only a small quantity oflubricant. Moreover, shown in FIG. 3 is an embodiment, in whichlubricant supply line 4 is connected to a pressure reservoir 14. Thepressure level of pressure reservoir 14 may advantageously be keptsubstantially constant between an upper and a lower limit value. To thisend, a pressure pump 13 is used, which is interposed in a controlcircuit for keeping the lubricant pressure constant.

Also shown in FIG. 1 is the arrangement of a metering device 15 servingfor the supply of an accurately metered quantity of lubricant in supplyline 4. Such a metering device may be, for example, an externallycontrollable valve, which is opened from time to time. Likewise, it isconceivable to use two channel sections rotating relative to oneanother, which overlap each other per shaft rotation one time in meshingengagement, so as to provide during the time of the overlap a continuouspassage from lubricant supply line 4 to outlet end 23 of duct 3.

As is still further shown in FIG. 1, a sealing disk 16 may be arrangedon each side of the respective rolling element zones, so as to preventon the one hand an uncontrolled discharge of the lubricant, for example,by evaporation. On the other hand, such sealing disks are useful, sincedepending on the installed situation, they avoid an unwanted entry ofdirt into the interior of the bearing. An unwanted entry of dirt wouldcounteract the desired lubrication effect.

Shown in FIG. 4 is a schematic diagram for the individual control of thelubricant supply to a plurality of bearing points.

From a machine control system 80, the basic adjustment resulting fromthe operating parameters is input via line 81 into a control unit 82.Input in control unit 82, on the other hand, are, via measuring lines83, operating data received from the individual antifriction bearings Iand output by measuring sensors 84, such as, for example, thetemperature values or the values originating from the measuring ofvibrations of the individual antifriction bearings. Control commandsresulting therefrom are supplied, via control lines 85 to individualmetering pumps 15, which again supply the respective antifrictionbearings 1, via lubricant lines 4, with individually determined amountsof lubricant.

To this end, lubricant is removed from a tank 47 and supplied via line86 to pumps 15, which, as shown, are interconnected via a collectingline 62.

The bearings 1 may be associated, for example, with a spinning machine.They may be part of one or more spindles driven at a high speed andcarrying winding tubes, or be associated with spindles which drive theformed packages or reciprocate the yarns. For example, they may also bethe bearings of feed rolls and/or draw roll of such a spinning machine.

FIG. 5 illustrates a device 2 for supplying a lubricant to antifrictionbearings 1. The antifriction bearings form part of a rotatable godet ona textile machine not shown in more detail for processing filamentyarns. It is a characteristic of this embodiment that a godet 45 isnonrotatably connected with a mandrel 46, which is supported in theinner rings 21 of antifriction bearings 1. For this reason, the outerrings 20 of the antifriction bearings 1 are installed nonrotatingly inthe housing. Suitably, the lubricant supply line 4 extends thereforefrom the nonrotating housing elements toward antifriction bearing 1. Itis a further characteristic of this embodiment that each lubricantsupply line 4 extends via a duct 3 to terminate at 23 in the nonrotatingouter ring 20 of the bearing in the region of the outer race of therolling elements.

The lubricant is delivered by means of a lubricant pump 13 from a tank47 into a closed-circuit supply line 48. To this end, the lubricant pump13 is arranged in the forward flow portion of a closed-circuit supplyline 48. The return flow portion of the closed-circuit supply lineterminates via a throttle 49 and a non-return valve 50 back in tank 47.The non-return valve 50 is biased by a compression spring, which loadsthe valve body from the side of the tank.

For the principle of this invention, a closable end of the return flowportion of the closed-circuit supply line would be sufficient (forexample, a controllable stop valve). A throttle in combination with thesubsequent non-return valve is not absolutely necessary, butadvantageous as regards the control.

Tap lines 39-41 branch off from the closed-circuit supply line 48 in theillustrated embodiment. A separate metering device 15 is provided foreach antifriction bearing 1. Likewise, it is conceivable that only asingle metering device 15 supplies the lubricant in measured quantitiesvia a corresponding line system leading to a plurality of antifrictionbearings 1.

As one can further note, the forward flow portion of closed-circuitsupply line 48 accommodates a pressure reservoir 14, which is acontainer filled with a lubricant 33, and in which a compressiblepressure medium, for example air, is compressed above the lubricantlevel. As shown, the pressure reservoir has no separate connection forthe gas, but one can presume that the lubricant displaces and therebycompresses the gas volume enclosed inside the pressure reservoir, sothat a portion of the energy applied within the gas cushion is stored,so as to subsequently expand stepwise as the lubricant is metered.

Further shown is a pressure relief valve 52, which is to be provided forsafety reasons, so as to open in response to a certain upper safetylimit value. Proceeding from lubricant pump 13 in the direction of theforward flow, a non-return valve 53 is provided which prevents thelubricant from flowing back when pump 13 is shut down.

Further provided along closed-circuit supply line 48 are the several taplines 39-41 which are followed by metering devices 15, and which, asaforesaid, are each supplied via closed-circuit supply line 48. Eachmetering device 15 is designed and constructed as an electricallycontrolled piston pump with a constant piston stroke. This permits asimple type of construction with an accurately predetermined deliveryvolume per piston stroke.

For a control of metering devices 15, which is common in thisembodiment, a control unit 28 not shown in greater detail is used whichis adjustable, if need be, with respect to control time and controlintervals.

Each piston pump is separated, via a further non-return valve 56 fromoutlet end 23 of the associated antifriction bearing. The additionalnon-return valve 56 opens, as shown, in direction toward antifrictionbearing 1.

As can further be noted, each tap line 39-41 extends from theclosed-circuit supply line 48 vertically downward. On principle, it willsuffice to lay the tap lines 39-41 such that they proceed from theclosed circuit supply line with a downgrade. Already in this instance,any as undissolved air bubbles would be forced to move, caused bybuoyancy, in a direction toward closed-circuit supply line 48 and wouldthen be transported in same, during the next flushing process, indirection toward tank 47. In this manner, an undersupply of lubricant tothe antifriction bearing with air-enriched lubricant is reliablyprevented.

Furthermore, each metering device 15 is provided with a pressure monitor54, which picks up the inlet pressure of metering device 15. Should thisinlet pressure fall below a certain lower limit value, for example, 1bar, the corresponding antifriction bearings or bearings will be shutdown, since a supply of lubricant is no longer guaranteed. If needarises, the shutdown may occur after a certain delay time, it beingpresumed that any lubricant still remaining in antifriction bearing 1cannot be suddenly used up.

Continuing further in the direction which coincides with the flushingdirection toward the tank, behind the last tap line 41 in closed-circuitsupply line 48, there is provided a throttle 49 which is followed bynon-return valve 50. In connection therewith, a pair of pressuremonitors is arranged in the forward flow portion, of which one monitorsan upper limit value 38, and the other a lower limit value 37. Uponreaching the upper limit value, for example, 3.8 bar, this pressuremonitor stops lubricant pump 13. At this time, the pressure inclosed-circuit supply line 48 amounts to 3.8 bar. On the other hand, theclosing pressure of non-return valve 50 (caused by the bias of thecompression spring), is less than 3.8 bar, for example, 3 bar. Thepreceding throttle 49 causes the pressure to drop, as long as thelubricant flows in the closed-circuit supply line, so that betweenthrottle 49 and biased non-return valve 50 the pressure is always lowerthan in the closed-circuit supply line. The higher pressure in theclosed-circuit supply line causes the lubricant to flow toward thereturn flow portion, at least as long as the pressure behind thethrottle is higher than the closing pressure of non-return valve 50.Upon reaching the closing pressure, non-return valve 50 closes as aresult of being biased by the compression spring, and any pressure stillpresent in closed-circuit supply line 48 remains stored.

As a result of continuously removing lubricant, this pressure declines,however, in the course of time. Upon reaching the lower limit value 37,the second pressure monitor, which is a closer, restarts lubricant pump13. The latter then pumps again the lubricant from tank 47 into theclosed-circuit supply line, until the first pressure monitor, which isan opening contact, disconnects the pump again. During this operation,the closed-circuit supply line is flushed. After the pump is shut down,the flushing operation continues until non-return valve 50 closes again.

Essential is here that the closing pressure of the subsequent non-returnvalve 50 lies between the upper limit value 38 and the lower limit value37, with the preceding throttle 49 causing a certain pressure drop, sothat the pressure of closed-circuit supply line 48, reduced by thepressure drop at the throttle, is present at the subsequent, biasednon-return valve 50.

In the place of the exclusively pressure-controlled non-return valve, itis also possible to provide an externally controllable stop valve, whichopened for flushing, while lubricant pump 13 is in operation, and isthen closed, preferably, while the lubricant pump is still slowing down.

Further provided is a float switch 55 for a continuous monitoring of thetank contents. Upon falling below a minimum level, a warning signal or ashutdown signal will be emitted, if need be.

The pumps 15 illustrated in FIGS. 6 and 7 show further details forcarrying out the invention. These pumps are suitable for metering verysmall amounts of fluid, in particular lubricating oil, for a meteredlubrication of antifriction bearings in accordance with this invention.The pump 15 may be rated such that it delivers per pump cycle (operatingstroke) no more than 3 mm³ of lubricant, it being possible to extend thedistribution of this small amount of fluid, via a corresponding controlof the pump piston, over more or less long periods of time.

The pump 15 is provided with a housing. Shown is only a portion 57 ofthe housing. Accommodated in the housing portion 57 is a pump chamber orcylinder 58, the volume of which determines the volume of thelubricating oil that is to be delivered per pump cycle (operatingstroke). The cylinder 58 communicates with, at its one end, a radialbore 59, which is hereafter termed inlet chamber, and which consists ofa closed portion 60 and an opposite portion 61 extending on the otherside of cylinder 58 and connecting to a fluid tank not shown. Theportion 61 is intersected by a collecting channel 62, which extendsthrough the housing and, if need be, merges into similar channels offurther pumps to be coupled with pump 15 (see FIG. 4).

Furthermore in portion 61, a line 63 terminates, which is connected witha pressure switch 54. The latter controls, regardless of the pressure inportion 61, a drive 65 for a pump piston 66 operating in cylinder 58, asis schematically shown, and under circumstances, also the drive of theelement, the bearing of which is lubricated by pump 15 (see FIG. 5). Thepump 15 is connected via a plug coupling 64 with a fluid tank orclosed-circuit supply line 48 (see FIG. 5).

The piston 66 slides in a guideway 67, which extends coaxially withcylinder 58 on the opposite side of inlet chamber 61. In the idleposition, the end surface of piston 66 extends into inlet chamber 59near the one end of cylinder 58 or the boundary edge between cylinder 58and inlet chamber 59. Thus, the piston 66 including the boundary edgeform the inlet valve of the pump.

Located at the end of cylinder 58, removed from inlet chamber 59, is theoutlet valve of pump 15, which is formed by a boundary edge surroundingthe outlet opening of cylinder 58, and a frustoconical piston 68. In theclosed state of the outlet valve, the tip of piston 68 is pressedagainst the boundary edge by a spiral spring 69 operative on theopposite end surface of piston 68. The boundary edge merges into aconical valve seat 70, which terminates in an outlet chamber 71. Thelatter accommodates spring 69 which biases piston 68. On the side ofinlet chamber 71, a line 4 starts which leads to a lubricating point notshown, for example, a godet bearing (see, FIG. 5).

The cone of piston 68 consists of a relatively soft material, forexample a polymer, while the boundary edge of the opening forming thevalve seat is made of a hard material. The fact that the vertex angle ofvalve seat 70 is greater than the vertex angle of piston 68 ensures onthe one hand a good sealing effect. On the other hand, the wedge-shapedannular gap which is formed by the different conicities improves thefluid flow through valve 15, when the outlet valve is opened, it beingassumed that air which is entrained in the lubricating oil and, undercircumstances, would collect on the boundary edge in the form of an airbubble, is discharged through this gap along with the oil. The sealingeffect of the outlet valve is further improved in that the soft materialof piston 68 adapts itself to the shape of the boundary edge.

The end of piston guideway 67 removed from inlet opening 59 connects viaan opening to a widened recess 73. The piston 66 which has the samediameter over its entire length, extends through this opening intorecess 73, and has at its end located therein a widened portion or head74. Operative on this head 74, is one end of a spiral spring 75 whichsurrounds piston 66, while its other end is supported in recess 73. Thespring 75 allows to keep piston 66 in its idle position, in which thefront end of the piston extends into inlet chamber 59 at a smalldistance from the boundary edge of cylinder 58.

Located below recess 73 and fixedly connected with the pump housing isdrive 65 for piston 66. The drive 65 is shown as an electromagnet, itbeing preferred to employ a so-called wet magnet, which is protected bya suitable enclosure against oil which may possibly leak from the pump,and which is periodically energized, so as to drive piston 66 intocylinder 58. Such electromagnets are known by their structure andoperation.

The advantage of a wet magnet layout consists in that no additionalseals are needed, and thus the risk of secondary air being mixed intothe lubricant is eliminated. As seen in FIG. 7, an additional floodchannel 77 may be provided between inlet chamber 61 and recess 73 whichserves as a rotor chamber, and overflow channels 78 which extend alongthe direction of movement of the rotor and through same, may be providedso as to realize during the movement of the rotor a fast overflow of thelubricating oil from the one front end of the rotor to its oppositefront end.

As aforesaid, the pump 15 as illustrated allows to deliver fluidquantities of about 3 mm³ per pump cycle or piston stroke. Prerequisitetherefore is, among other things, that the dimensions of both cylinder58 and piston 66, and the stroke transmitted on same by drive 65 areaccordingly small. Thus, the piston diameter may measure about 2 mmwide, and the piston stroke about 1 mm long. The front end of piston 66lies in its idle position a very slight distance away from the cylinderinlet, or the boundary edge, and in its advanced position, the pistonend surface comes to lie at a short distance from the front end ofpiston 68 directed against cylinder 58, without the two contacting eachother. The oil pressure developing as a result of the upward movement ofpiston 66, not only causes the valve piston 68 to rise and the outletvalve to open, but also pumps out air which is entrained by thelubricating oil. This avoids that an air bubble forms, which would leadin a pump with such a short delivery stroke to considerable disturbancesin its pumping operation.

The pump of FIG. 7, with small exceptions which will each be identified,corresponds to the embodiment of FIG. 6 as described above. Moreparticularly, the pump of FIG. 7 comprises a housing which is composedof a portion 57 and a housing for drive 65. Both housing portions areinterconnected in airtight manner. Formed in housing 57 is an inletchamber which has the shape of a blind-end bore with sections 59 and 61.The open end of this blind bore is connected with an inlet channel. Theconnection is made by a plug coupling 64 with a non-return valve. Whenjoining the coupling elements, the non-return valve is opened by aplunger. When separating the coupling elements the valve closes, so thatthe oil is unable to escape from the supply line.

Between bore sections 59 and 61, the inlet chamber is verticallyintersected by a further bore with portions 58 and 67. This bore extendsthrough the inlet chamber at a distance from the blind hole end of theinlet chamber, thereby dividing the blind-end bore into the portion 59in the region of the blind hole end, and the inlet portion 61. The borevertically intersecting the inlet chamber forms with its one branch 58 acylinder, which terminates on the side facing away from the inletchamber in an outlet chamber 71. In its inlet end region, the outletchamber 71 is provided with a conical valve seat, which extendscoaxially with cylinder 58. Arranged in the conical seat is a conicalvalve body 68, which is pushed by a compressing spring into the valveseat 70. The angle of cone of valve body 68 is smaller than the angle ofcone of valve seat 70. The tip of valve body 68 is cut off, so that thesmall diameter of valve body 68 corresponds substantially to thediameter of cylinder 58. Therefore, as will be described below in moredetail, the small conical surface of valve body 68 is suitable to serveas a stop for the pump piston. The outlet chamber 71 has an outletchannel 4 which connects to a lubricant supply line. This lubricantsupply line leads, for example and in particular, to the hole in one ofthe above described antifriction bearings.

Portion 67 of the bore intersecting the inlet chamber, which faces awayfrom cylinder 58, serves as a guideway for piston 66. This piston 66 isconstructed as a circular-cylindrical pin. The diameter of this pin isadapted within close tolerances to the diameter of cylinder chamber 58.The embodiment of FIG. 6 has the characteristic that the portion of thebore serving a guideway 67 is larger than the cross section of the pin,thereby permitting the oil to flow from the inlet chamber into recess 73which will be described below. In the embodiment of FIG. 7, the portionserving as guideway 67 may be within closer tolerances, since in thisembodiment, an additional flood channel 77 extends from the inletchamber and connects the latter with recess 73.

The recess 73 is a circular-cylindrical bore which is arranged,coaxially with the bore serving as cylinder 58 and guideway 67 in theone side of housing portion 57, namely, on the side of the bore sectionserving as guideway 67, which faces away from the inlet chamber. Thepump piston 66 measures so long that it is able to move with its onefront end between the inlet chamber and the small conical surface ofvalve body 68. During this movement, the piston 66 extends with itsother end facing away from the inlet chamber, into recess 73. At thisend, the piston is provided with a shoulder (head 74). Supported on thisshoulder, on the one hand, and on the opposite end surface facing theinlet chamber, is a compression spring 75. This compression springforces the piston 66 into an idle position, in which its front end(control end surface 87) facing cylinder 58 lies in the inlet chamber,i.e., it does not close the edge of penetration between cylinder 58 andbore 59, 61 forming the inlet chamber, hereafter referred to as controledge 88. This control edge 88 forms together with control end surface 87the inlet valve of cylinder 58.

On its free side, recess 73 is sealed against fluid by the housing ofdrive 65. The drive is an electromagnet comprising an iron plunger(rotor 89) and a toroidal coil 90. The latter is embedded in the housingin fluidtight manner and connected, via lines not shown, with a controldevice. The rotor is straight guided in a chamber 91, which is realizedby two guide bores 92 and 93. The guide bore 92 is a blind hole which isformed in the side of rotor chamber 91, which faces away from the otherhousing portion 57. On this side, the rotor 89 is provided with a guidepin 94, which slides in guide bore 92. The cross section of guide pin 94is substantially smaller than the cross section of guide bore 92. Thisallows oil to penetrate from rotor chamber 91 into guide bore 92. Theguide bore 93 connects on the one hand rotor chamber 91 with recess 73,and is used on the other hand for guidance. To permit an unimpeded oilpassage, the cross section of an actuation plunger 95, which is attachedto the rotor, is smaller than the cross section of guide bore 93.

It should be emphasized that rotor 89 and actuation plunger 95 arearranged along the axis of piston 66. The rotor 89 is provided with axisparallel overflow channels 78, which interconnect the two ends of rotorchamber 91. The actuation plunger 95 cooperates with the end of piston66 facing away from control end surface 87. In the nonenergized state ofring magnet coil 90, spring 75 forces the piston, the actuation plunger95, as well as rotor 89 into an end position, in which, as previouslydescribed, the control end surface 87 of piston 66 extends into theinlet chamber, and does not close control edge 88 of cylinder 58. Uponenergizing ring magnet coil 90, rotor 89, actuation plunger 95, andpiston 66 are displaced so far that control end surface 87 initiallycloses control edge 88, subsequently immerses into cylinder 58, forcesthe fluid from cylinder 58 against the pressure of valve spring 69 byopening valve body 68, and finally also contacts valve body 68.

Referring to the foregoing description, it should be remarked that itwill suffice, when the control end surface 87 comes to a stop at a shortdistance from the valve body 68 being seated in its seat. This will bepermissible in particular, when the lubricant oil does not enclose anylarger amounts of air. However, should air collect in cylinder 58, therisk may arise that residual air is not entirely forced out. In thisevent, it will be useful that end surface 87 advances so far that valvebody 68 is unable to close entirely under the pressure of spring 69.This allows to ensure that also residual air is able to escape.

It should here be expressly emphasized that the oil is supplied to theinlet chamber under pressure. It is not the primary function of theabove described metering pump to increase pressure of the oil or performthe intake itself.

The characteristic of the pump and its here described construction,which also applies to the embodiment of FIG. 6, consists however in thatthe pump does not cause any pressure fluctuations whatsoever of thelubricant in the inlet chamber, in particular the oil, and thus likewiseno pressure fluctuations during the filling of cylinder 58. To thisextent, the pump differs from all known pumps, in which the pulsation ofthe fluid flow on the outlet side is accompanied by a correspondingpulsation on the inlet side. In the described pump, neither the pistonmovement, nor the movement of the magnet rotor, nor the actuationplunger, nor the guide plunger lead to a change of the total volume,which is contained in the inlet chamber, in recess 73, and in rotorchamber 91, although the delivery is discontinuous.

In this context, it should also be noted that it is possible to joinseveral pump housings by flanging. In this event, it will be opportuneto provide only one inlet channel for all interconnected pumps. Thepumps will then be joined together by a collecting channel 62, whichintersects vertically the bore of all inlet chambers of theinterconnected pumps.

Likewise, in this event, a single pressure monitor 54 will suffice,which performs a desired switching, when the pressure drops in the inletchamber, for example, when it falls below a minimum value and themachine shuts down, so as to prevent a dry operation. In this instance,each of the pumps will serve as a metering pump for an antifrictionbearing.

In the operation of the pumps, lubricating oil is supplied under acertain pressure, for example 2 bar, via the radial bore or inletchamber 59, to cylinder 58. In the illustrated, idle position of piston66, the lubricating oil flows not only into cylinder 58, but also intothe closed portion 60 of inlet opening 59. As a result, piston 66 isbiased by centripetally the same pressure, which is highly significant,especially in view of the small diameter of piston 66. In an upwardmovement which is triggered by the magnet, piston 66 enters into thecylinder, whereby the oil therein flows into outlet chamber 71 byraising valve piston 68, and subsequently, via line 4 to the point to belubricated. Thereafter, when piston 66 is returned by the force ofspring 75 to its illustrated position, piston 68 of the outlet valvecloses the outlet opening, and a vacuum forms in cylinder 58 as a resultof the downward movement of piston 66. As soon as the end surface ofpiston 66 emerges from cylinder 58 and slides into inlet opening 59,i.e. toward the end of the intake cycle of piston 66, the vacuum causeslubricating oil to refill cylinder 58 very rapidly and independently ofthe movement of piston 66. Thereafter, the pump is able to perform afurther pumping cycle.

Each of FIGS. 8, 8a, and 9 illustrate an antifriction bearing 1 with adevice 2 for supplying a lubricant. This device 2 comprises a lubricantsupply line 4, through which the lubricant is delivered to antifrictionbearing 1. In the region of antifriction bearing 1, lubricant supplyline 4 has an outlet end 23 from which the lubricant exits in directiontoward antifriction bearing 1. This special case includes in additionthe characteristic that the lubricant supply line 4 extends through theouter ring 20 of the antifriction bearing, and that the lubricant supplyline 4 terminates substantially in the region of the outer or inner race6a or 6i of the rolling elements in direction of the rolling elements 5.

It is, however, expressly stated that this invention it not limited tothis structural form. In particular, the lubricant supply line mayextend also from a lateral direction to rolling elements 5, and bedirected in the region of the cage of the rolling elements laterally indirection toward the rolling elements.

In the present embodiments, the lubricant is put under an increasedpressure in lubricant supply line 4. As is indicated by pressure gauge27, this pressure is understood to be above the atmospheric pressure. Inthe end region of lubricant supply line 4, a controllable valve 15 isarranged, which retains the lubricant put under increased pressure inline 4, as long as this valve 15 is closed. As can be noted, the valvecan be actuated by a control device 28. As need arises, the latter opensand closes the valve. The lubricant being thereby caused to exit underincreased pressure, leaves outlet end 23 of lubricant supply line 4, aslong as valve 15 is opened. Upon closing valve 15 by means of controldevice 28, the exiting lubricant flow is separated, and the lubricationperiod is completed.

A characteristic of this invention lies in that the control devicepossesses two possibilities of adjustment. To this end, it is possibleto adjust by means of a first adjusting device 29 the duration ΔT, forexample, every two hours. This control device thus allows to meter thequantity of lubricant exiting each time, as a function of the durationof its exit, as well as the time of its exit, in the most accuratemanner, so that only the quantity of lubricant is supplied which isrequired by antifriction bearing 1.

To this end, as can be noted, the control device 28 is operative on theelectric control of the valve. In the present embodiment, the valve isactuated by an electrically controllable magnet 31.

As best seen in FIG. 8, the existing lubricant supply lines can easilybe developed into pressure reservoirs, it being preferred to close theselubricant supply lines 4 toward the inlet side by a non-return valve 32.The function of this non-return valve is that the pressure built-up inlubricant supply line 4 remains stored free of losses.

The length of lubricant supply line 4 may, if desired, be enlarged bythe installation of additional loops, if a certain storage volume is tobe realized.

As shown in FIG. 8a, it is also possible to provide a closed-circuitsupply system 48, which may be structured in accordance with FIG. 5.

As is shown moreover in FIG. 9, the lubricant can be supplied from apressure reservoir 14, which may be biased by a pressure medium,preferably air or another gas. To this end, a special pressure reservoir14 is provided, which contains on the one hand a lubricant 33. Above thelubricant level, pressure reservoir 14 is biased by a pressure mediumsupplied via a line 34. To this end lubricant pump 13 delivers inpressure line 34 only the medium for increasing the pressure, in that ittakes in ambient air and, if need be, after a corresponding filteringand drying, pumps same into pressure reservoir 14. As a result, a volumeforms in the pressure reservoir above the lubricant level, which causesthat lubricant 33 is supplied under pressure to the lubrication pointsof antifriction bearing 1, even in the absence of the respective pumpoperation.

As can further be noted, the pressure reservoir 14 is provided with apressure gauge 27. In this embodiment, the latter acts as a start-stopswitch, which connects or disconnects lubricant pump 13, via a switchingdevice 36. Once the existing pressure level falls below the lower limitvalue 37, pressure gauge 27 starts lubricant pump 13, via switchingdevice 36, for example, by means of a relay. As a result, the pressureincreases in pressure reservoir 14. As soon as the pressure gaugeregisters the reaching of the upper limit value 38, lubricant pump 13 isdisconnected by switching device 36, until the pressure of pressurereservoir 14 reaches again its lower limit value. Further shown in FIG.9 is in phantom lines that lubricant supply line 4 includes a pluralityof tap lines 39-41 for supplying lubricant to several antifrictionbearings 1. As can be noted, in this instance, the controllable valve 15can be arranged in the common lubricant supply line 4 of all tap lines39-41. In this instance, the advantage presents itself that, with littleconstructional expenditure, all tap lines are jointly controlled.

However, if far removed antifriction bearings are to be supplied withlubricant, it will be opportune to employ the embodiment shown inphantom lines. In this instance, a corresponding valve 15 is arranged atthe end of each tap line. All these valves 15 may be actuated jointly atthe same time, or individually in accordance with the lubricantrequirements determined for each antifriction bearing.

This has the advantage of yet a cost-favorable construction, since onlya single control device will be needed, which actuates in parallelconnection all metering valves at the same time, or drives each time viacarrier frequencies only certain metering devices.

Shown in FIG. 10 is the invention put to practice in a textile machinefor producing endless, synthetic filament yarns.

A yarn 101 of a thermoplastic material is spun. The thermoplasticmaterial is supplied by means of a feed hopper 102 to an extruder 103.The extruder 103 is driven by a motor 104 which is controlled by a motorcontrol unit 149. In extruder 103, the thermoplastic material is melted.To this end, use is made of the deformation work (shearing energy) whichis applied to the material by the extruder. In addition, a heatingsystem 105, for example, in the form of a resistance heater is provided,which is activated by a heater control unit 150. Through a melt line106, in which a pressure sensor 107 is provided for measuring thepressure of the melt intended for controlling the pressure and speed ofthe extruder, the melt reaches a gear pump 109, which is driven by apump motor 144. The latter is controlled by a pump control unit 145 suchas to permit to very finely adjust the pump speed. Pump 109 delivers themelt flow to a heated spin box 110, the underside of which accommodatesa spinneret 111, from which the melt exits in the form of sheets of finefilaments 112. The latter advance through a cooling shaft 114, in whichan air current is directed transversely or radially to the sheet offilaments 112, thereby cooling the filaments.

At the end of cooling shaft 114, the sheet of filaments is combined byan spin finish roll 113 to a yarn 101 and provided with a finish. Fromcooling shaft 114 and spinneret 111, the yarn is withdrawn by a deliveryroll or godet 116, about which the yarn loops several times. To thisend, a guide roll 117 is used, which is arranged offset with respect togodet 116. The latter is driven by a motor 118 and frequency converter122 with preadjustable speed. This delivery speed is by a multiplehigher than the natural exit speed of filaments 112 from spinneret 111.

Arranged downstream of godet 116 is draw roll or godet 119 with afurther guide roll 120. Both correspond in their arrangement to deliveryroll 116 with guide roll 117. For the drive of draw roll 119, a motor121 and frequency converter 123 are used. The input frequency offrequency converters 122 and 123 is evenly predetermined by acontrollable frequency generator 124. This allows to individually adjuston frequency converters 122 and 123 the speed of delivery roll 116 anddraw roll 119 respectively. The speed level of delivery roll 116 anddraw roll 119 is however collectively adjusted on frequency converter124.

From draw roll 119, the yarn 101 advances to a so-called "apex yarnguide" 125 and thence to a traversing triangle 126, where a known yarntraversing mechanism 127 (not shown) is provided. The latter comprises,for example, oppositely rotating blades, which reciprocate the yarn 101over the length of a package 133. In so doing, downstream of traversingmechanism 127, yarn 101 loops about a contact roll 128, which restsagainst the surface of yarn package 133. The latter is wound on a tube135, which is mounted on a winding spindle 134. Winding spindle 134 isdriven by a motor 136 and spindle control 137 such that the surfacespeed of package 133 remains constant. To this end, the speed of freelyrotatable contact roll 128 is sensed as a control variable on contactroll shaft 129 by means of a ferromagnetic insert 130 and a magneticpulse generator 131.

Analogously, the foregoing also applies to both a ferromagnetic insert138 and pulse generator 139 of spindle 134.

It should be mentioned that yarn traversing system 127 may also be astandard cross-spiralled roll with a traversing yarn guide reciprocatingin a cross-spiralled groove across the traversing range.

The textile machine, of which only one processing station is shown,comprises a plurality of individual antifriction bearings 1, which arestructured corresponding to the above described antifriction bearings 1and supplied with a metered amount of lubricant. Their foregoingdescription is herewith incorporated by reference.

Illustrated in FIG. 11 are three possible embodiments of an antifrictionbearing, in which the outlet end 23 of duct 3 can also be arranged onthe side of the rolling element race, which is biased by the transverseforce 11 of the bearing. To avoid in this embodiment that the rotatingrolling elements roll over outlet end 23, thereby closing it in thecourse of time, the race is provided with at least one, preferablyhowever several, annular rolling zones 150, which permit an accuratelydefined rolling motion of rolling elements 5.

Such antifriction bearings are described as so-called multipoint ballbearings, with the number of points indicating the number of annularrolling zones, which the rolling element contacts with respect to bothraces. Preferably, the geometrical arrangement of annular rolling zonesis such that the rolling condition is met in each point of the annularrolling zones.

The first embodiment is illustrated in FIG. 11a and comprises aso-called four-point ball bearing. Formed in the outer and the innerring are each two annular rolling zones 150. Therebetween, contactfreeannular zones 151 are formed, which are described within the scope ofthe present invention as the no load zone in the race of the rollingelements. In the illustrated four-point ball bearing, the no load zoneof the race extends between the two annular rolling zones 150 arrangedin each bearing ring. In this zone, there is no contact between thebearing ball and its associated race. Consequently, it is natural toarrange outlet end 23 of hole 3 in this region. By way of example, it isassumed that in the present embodiment hole 3 is arranged in the outerbearing ring.

Contrary thereto, the second embodiment, as shown in FIG. 11billustrates a so-called two-point ball bearing, in which, related to anaxial section, the curvature of the ball is greater than the curvatureof its respective race 6a, 6i. As a result, only one annular rollingzone is created in each race, while a no load zone of the race is formedon the side of each annular rolling zone.

Characteristic of this zone is a slight, radial spacing between theball, as it passes by, and the surface of the race. Generally, in thisspace, there is no contact between the ball and the surface of the race.An occasional contact however is harmless. It is therefore possible toarrange in this zone of the race the outlet end 23 of duct 3 also on thebearing side which is biased by the transverse force of the bearing.

Moreover, within the scope of this invention, it is still furtherpossible to employ a so-called three-point ball bearing, as is shown inthe illustration of FIG. 11c. The foregoing description applies to thisillustration in corresponding manner.

Supplementing the foregoing, shown in FIG. 12 is an embodiment of thisinvention employing axially biased grooved ball bearings mounted on theshaft. They are two single row ball bearings which are mounted relativeto one another in the axial direction of the shaft. This is commonpractice in mechanical engineering, so as to keep the axial play of thesupported shaft as small as possible. For this reason, the left ballbearing rests with its outer ring against the offset portion of ahousing 96, thus defining clearly the position of the outer ringrelative to the housing. The end surface of the left ball bearing ispressed against the offset housing portion by means of a clamping cover97. To this end, the clamping cover 97 engages on the outer ring of theright-hand ball bearing and pushes same in direction toward the ballbearing on the offset housing portion.

This axially biased bearing is quite common, so that with respect to isdetails not shown, reference may be made to the state of the art.Essential however is that between the two ball bearings an axial stay isproduced, whereby the bearing balls are shifted in their respectiveraces such that the annular contact zones 150 do no longer lie in oneradial plane, but in an oblique plane slightly inclined thereto. Thecontact zones are formed as substantially annular rolling zones 150,which, when related to each ball bearing, are arranged in the onebearing ring to the right, and in the other bearing ring to the left ofthe central radial plane of the bearing. Thus, the annular rolling zoneundergoes a certain lateral displacement from the central radial planeof the bearing. As a result, the central radial plane of the bearing isrelieved from a load, so that a substantially contactfree, annular zoneis formed, which allows to accommodate the outlet end 23 of duct 3.

Consequently, in the case of axially biased bearings, it is possible toemploy the invention likewise with the use of a standard, single rowgrooved ball bearing, with an essential, additional advantage consistingin that the outlet end 23 of duct 3 can also be arranged in the centralradial plane of the bearing, inasmuch as it is a substantiallycontactfree annular zone, which forms the no load zone of the ball race.

What is claimed is:
 1. A method for a controlled supply of lubricant toan antifriction bearing which comprisesan inner bearing ring and anouter bearing ring mounted in a concentric arrangement so as to defineopposing inner and outer races on the inner and outer ringsrespectively, a plurality of rolling elements mounted between the innerand outer races, and a duct extending through at least one of the innerand outer rings and terminating at an outlet opening which communicateswith the associated race, said method comprising the steps ofintermittently supplying lubricant to the duct of the antifrictionbearing so as to flow through the outlet opening and directly into theregion of the rolling elements, with the amount of the lubricant foreach intermittent supply being determined from a predetermined basicquantity which is modified as a function of operating data which arecontinuously acquired from the antifriction bearing.
 2. The method asdefined in claim 1 wherein the operating data comprise the temperatureof the antifriction bearing.
 3. The method as defined in claim 1 whereinthe operating data comprise the vibration of the antifriction bearing.4. The method as defined in claim 1 wherein the operating data comprisea vibration selected from a defined range of frequency.
 5. The method asdefined in claim 1 wherein the operating data comprise the amplitude ofthe vibration of the antifriction bearing.
 6. The method as defined inclaim 1 wherein the operating data comprise the frequency of thevibration of the antifriction bearing having a selected amplitudeheight.
 7. The method as defined in claim 6 wherein the frequency isselected within a range between 200 kHz and 500 kHz.
 8. The method asdefined in claim 1 wherein the lubricant is intermittently supplied tothe duct of the antifriction bearing by means of a pump which deliversthe lubricant under an increased pressure and in a metered quantity. 9.The method as defined in claim 1 wherein the lubricant is intermittentlysupplied to the duct of the antifriction bearing via a pressurizedlubricant reservoir and a valve connected to the reservoir which isintermittently actuated.
 10. The method as defined in claim 1 whereinthe step of intermittently supplying lubricant to the duct of theantifriction bearing occurs via a sequence of tripping signals of apredetermined frequency and controlled duration.
 11. The method asdefined in claim 1 wherein the step of intermittently supplyinglubricant to the duct of the antifriction bearing occurs via a sequenceof tripping signals of a predetermined duration and controlledfrequency.
 12. A method for a controlled supply of lubricant to each ofa plurality of antifriction bearings in a multiposition textile yarnprocessing machine, with each of the antifriction bearings comprisinganinner bearing ring and an outer bearing ring mounted in a concentricarrangement so as to define opposing inner and outer races on the innerand outer rings respectively, a plurality of rolling elements mountedbetween the inner and outer races, and a duct extending through at leastone of the inner and outer rings and terminating at an outlet openingwhich communicates with the associated race, said method comprising thesteps of intermittently supplying lubricant to the duct of eachantifriction bearing so as to flow through the outlet opening anddirectly into the region of the rolling elements, with the amount of thelubricant for each intermittent supply being determined from apredetermined basic quantity which is modified as a function ofoperating data which are continuously acquired from the antifrictionbearing.